5-200

How often have you heard the statement “it isn’t cooling?”Well it may seem a bit picky, but it is not entirely accurate tosay the refrigeration system “cools”. If the system is operatingproperly, the refrigerated space should be “cooler” than its sur-roundings, but it is the result of a heat transfer process. Heat istransferred from the refrigerated space to the refrigerant, andultimately from the refrigerant to the ambient (at the condens-er). A lower temperature in the refrigerated space is thebyproduct of this heat transfer process. Perhaps this is a minorshift in thinking, but in viewing the refrigeration system forwhat it is…a heat transfer process…a more fundamentalapproach for diagnosis may be obtained.

In an effort to gain a better understanding of the various heattransfer processes occurring in a refrigeration system, the pres-sure-enthalpy chart can be of great use. Additionally, onceunderstood, “the chart” can be a tremendous benefit in analyz-ing the relative health of a refrigeration system. Let’s followthe refrigerant on a quick journey through a refrigeration sys-tem to see what it experiences, and plot it on “the chart” as wego. Before we start, a few technical definitions are in order:

Refrigeration - The achievement of a temperature below thatof the immediate surroundings.

Latent Heat of Fusion - The quantity of heat (Btu/lb) requiredto change 1 lb. of material from the solid phase into the liquidphase.

Latent Heat of Vaporization - The quantity of heat (Btu/lb)required to change 1 lb. of material from the liquid phase intothe vapor phase.

Sensible Heat - Heat that is absorbed/rejected by a material,resulting in a change of temperature.

Latent Heat - Heat that is absorbed/rejected by a materialresulting in a change of physical state (occurring at constanttemperature).

Saturation Temperature - That temperature at which a liquidstarts to boil (or vapor starts to condense). The saturation tem-perature (boiling temperature) is constant at a given pressure,*and increases as the pressure increases. A liquid cannot beraised above its saturation temperature. Whenever the refriger-ant is present in two states (liquid and vapor) the refrigerantmixture will be at the saturation temperature.

Superheat - At a given pressure, the difference between avapor’s temperature and its saturation temperature.

Subcooling - At a given pressure, the difference between a liq-uid’s temperature and its saturation temperature.

Ton of Refrigeration - The amount of cooling required tochange (freeze) 1 ton of water at 32ºF into ice at 32ºF, in a 24hour period.

Btu - British Thermal Unit: The amount of heat required toraise 1 lb. of water 1ºF.

1 Ton - 12,000 Btu/hr

Fig. 1 illustrates some of these definitions, using water as themedium experiencing a heat transfer process. This graph plotsthe water temperature vs. the enthalpy of the water (heat con-tent in Btu/lb). We all know that water boils at 212ºF (atmos-pheric pressure at sea level). By definition, water at atmospher-ic pressure, at a temperature lower than 212ºF, is subcooled.So, we start with subcooled water at 42ºF, and begin transfer-ring heat to it. Assuming we are working with 1 lb. of water,for every Btu added, a corresponding temperature increase of1ºF will be achieved (the definition for one Btu). It we contin-ue to add heat, eventually the water’s temperature will increaseto 212ºF (the saturation temperature at atmospheric pressure).At this point, the water begins to change states from a liquid toa vapor (boil). As noted on the graph, the water will experienceno further temperature increase…for a given pressure, thesaturation (boiling) temperature is the highest temperaturea liquid can ever achieve. Increasing the amount of heat trans-ferred to the water simply increases the rate at which the waterboils. If the temperature of the vapor were to be measured,we’d find it to be 212ºF (saturated vapor). Once the vapor hasseparated from the liquid, additional heat transferred to it willresult in a temperature increase. By definition, the vapor at232ºF (20º above the saturation temperature), is superheated.

It is interesting to note that while it takes only 1 Btu to raise 1lb. of water 1ºF, it takes almost 1000 times that amount (966.6Btu) for the 1 lb. mass of water to change states from liquid tovapor. A boiling liquid will always absorb more heat than avapor experiencing a temperature increase (per unit of mass).Understanding this principle explains why the evaporator in arefrigeration system should always be nearly filled with liquidrefrigerant. Otherwise, its full potential as a heat transferdevice will not be utilized.

The pressure-enthalpy chart, as shown in Fig. 2, displays allthe pertinent properties for a given refrigerant (in this exampleR22). The bubble to the left is the portion of the diagram wherethe refrigerant is in the saturated condition. The blue line onthe left of the bubble represents the 100% saturated liquid line,the thin dashed line on the right represents the 100% saturatedvapor line, and anywhere inside the bubble represents therefrigerant as a mixture of saturated liquid and saturated vapor.

To the left of the saturated liquid line is the area where therefrigerant can exist at a temperature lower than the saturated

The Pressure - Enthalpy ChartBy Dave Demma, Senior Application Engineer - Supermarket Refrigeration

January 2005 / FORM 5-200

* Except for zeotrope refrigerants

condition; subcooled liquid. To the right of the saturated vaporline is the area where the refrigerant can exist at a temperaturehigher than the saturated condition; superheated vapor. Thecritical point is the highest temperature that the refrigerant canexperience, and remain in the liquid form. If the temperatureexceeds the critical point, regardless of pressure, the refriger-ant can only exist in the vapor state.

All of the relevant properties are shown in Fig. 2:

Pressure - The vertical axis of the chart, in psia (see pink line).To obtain gauge pressure, subtract atmospheric pressure.

Enthalpy - The horizontal axis of the chart shows the heat con-tent of the refrigerant in Btu/lb.

Temperature - Constant temperature lines generally run in avertical direction in the superheated vapor & sub-cooled liquidportion of the chart. In the saturated bubble, the constant tem-perature line is along the horizontal, illustrating that the satura-tion temperature is constant at a given pressure (see black line).

Specific Volume - Constant volume lines extend from the redline saturated vapor line out into the superheated vapor-por-tion of the chart at a slight angle from the horizontal axis.

Specific volume is expressed in cu.ft/lb. (see orange line).

Entropy - Entropy is the mathematical relationship betweenheat and temperature, and relates to the availability of energy.These lines extend at an angle from the saturated vapor line.Their presence on the chart is relevant in that vapor compres-sion (in the ideal cycle) occurs at constant entropy (see darkblue line).

Quality - Lines of constant quality appear vertically, and onlywithin the saturation bubble. The refrigerant within the bubbleis a mixture of liquid and vapor at saturation, and the quality isthe percentage of the mixture which is in the vapor state (seegreen line).

The ultimate goal of the refrigeration system is to get therefrigerant into a condition where it can be useful as a mediumto transfer heat from the refrigerated space. If the systemdesign requires a -10ºF space temperature, you would expectthe refrigerant temperature in the evaporator to be somewhatlower…say -20ºF (a 10º temperature difference). This allowsthe relatively warmer air (something above the design of -10ºF)to be blown across the evaporator, flowing relatively coolerrefrigerant (-20ºF). The result is the transfer of heat from thewarmer air to the cooler refrigerant.

/ FORM 5-200

Water at Atmospheric Pressure(14.7 psia pressure)

Refrigeration

Boiling at constant pressure and temperature

SaturationTemperature

SuperheatedVapor

SubcooledLiquid

Latent Heat of Vaporization966.6 Btu per lb

20° Superheat (232°– 212°= 20°)

Enthalpy (Btu per lb)

Tem

pera

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(°F)

600 800 1000 1200400200

32

212

232

Fig. 1 – Water undergoing a change of state.

The purpose of the compressor is to take a low pressure vaporand compress it into a high pressure vapor. This occurs(in theory) at a constant entropy. In an ideal cycle, the refriger-ant vapor would enter the compressor as a saturated vapor…nosuperheat.

The Ideal CycleIf the operating temperatures and pressures are known, therefrigeration system can be plotted on the P-H diagram. Let’sassume our system is operating at a -20ºF evaporator and 100ºFcondensing.

The saturated vapor entering the compressor suction would beat -20ºF, illustrated by point #1 in Fig. 3. The vapor is com-pressed, following the constant entropy line to the pressure cor-responding to 100ºF, or 210.7 psia (point #2). The refrigerantvapor experiences a sensible heat gain during the mechanicalcompression process, resulting in a superheated vapor. This isillustrated by the location of point #2, to the right of the saturat-ed vapor line. In the ideal system, which does not considerpressure loss in the valves, refrigerant tubing, etc., point #2 rep-resents the outlet of the compressor/inlet of the condenser.

In the ideal cycle, the condenser serves as a two-fold compo-nent. Before any condensation occurs, the high pressure vapormust first be brought to a saturated condition (de-superheated).

Enough heat must be transferred from the refrigerant to lowerits temperature from 180ºF to the saturation temperature of100ºF (point #2A on the chart). At this point, condensation canbegin. As heat continues to be transferred from the refrigerantvapor to the air (or water, if a water cooled condenser is used),the quality of the refrigerant (% of the refrigerant in the vaporstate) will continue to decrease, until the refrigerant has beencompletely condensed. In the ideal system, this occurs at theoutlet of the condenser (point #3 on the chart). In the realworld, some subcooling would be expected at the condenseroutlet. Subcooled liquid provides insurance against liquidflashing as the refrigerant experiences pressure losses in thetubing and components.

The refrigerant is in the liquid state now, and at a high pressureand temperature. It must undergo one more change before itbecomes a useful heat transfer medium; a reduction in temper-ature. This is accomplished by reducing the pressure. You cancount on the refrigerant’s pressure-temperature relationship tobe an infallible law. If the pressure of a saturated liquid isreduced, the law governing its existence requires it to assumethe saturation temperature at the new pressure.

So, in order to reduce the temperature, the pressure has to bereduced, and some sort of restriction is required for this tooccur. It would be preferable if the restriction could regulate

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REFRIGERANT 22

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Fig. 2 – R22 Pressure-Enthalpy Diagram

FORM 5-200 /

itself as the system load demands change. This is exactly whatthe thermostatic expansion valve does; it is an adjustablerestriction which causes a reduction in liquid refrigerant pres-sure, yet will modulate in an effort to maintain constant super-heat at the evaporator outlet. The TEV is a superheat control,and will not maintain a constant evaporator pressure. It onlyprovides the restriction necessary to reduce the pressure tosome level, which will be determined by compressor size, TEVsize, load demand, and system conditions. If a constant evapo-rator temperature is required, it can be achieved very simply bymaintaining the pressure corresponding to the saturation tem-perature required. This is accomplished by adding an evapora-tor pressure regulating valve to the system.

Our ideal cycle has experienced a pressure drop in the TEV,and for the purpose of discussion we are at a constant 24.9 psiain the evaporator. This is a saturation temperature of -20ºF. Youwill notice that the refrigerant quality at the TEV inlet was 0%,and has increased to 35% at the evaporator inlet. Subcooling orsuperheat cannot exist where there is a mixture of liquid andvapor. Therefore, any place in the system where the refrigerantexists in two states (the receiver, parts of the evaporator andcondenser, the accumulator at times), it will be at the saturationtemperature for its pressure. For example, R22 in a saturatedstate at 24.9 psia (10.2 psig) will always be at -20ºF.

Because the refrigerant must conform to its law of existence,when the 210.7 psia (100ºF) liquid experiences the reduction inpressure to 24.9 psia, it must drop to the new saturation tem-perature of -20ºF. Some of the liquid refrigerant is required toboil as a means of removing the heat necessary to achieve thislower temperature. Yet another heat transfer process, whichyields a lower liquid temperature. The liquid that is sacrificedin the boiling process explains the increase in refrigerant qual-ity. The greater the difference between the liquid temperatureand evaporator temperature, the more liquid will have to beboiled to achieve the new saturation temperature. This resultsin an even higher refrigerant quality.

The final portion of the refrigerant’s journey is as a mixture ofsaturated liquid and vapor, traveling though the evaporator tub-ing. Warm air is blown across the evaporator, where its heatcontent is transferred to the boiling refrigerant. This is a latentheat gain to the refrigerant, causing no temperature increase,while experiencing a change of state. In the ideal cycle, the lastmolecule of saturated liquid boils off at the evaporator outlet,which is connected to the compressor inlet. Hence, the vapor atthe inlet of the compressor is saturated.

The cycle continues this way until the refrigerated space tem-perature is satisfied, and the equipment cycles off.

/ FORM 5-200

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IDEAL CYCLEIDEAL CYCLE

44

Condensation LineCondensation Line

Evaporation LineEvaporation Line

Compression LineCompression Line

Desuperheating LineDesuperheating Line

Expansion LineExpansion Line

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33 2A2A 22

Fig.3 - Plotting the Ideal Cycle on the P-H Diagram

Once the system has been plotted, various data points can beread and used for the system design calculations. Admittedly,this information is not typically something the technician willneed for servicing the refrigeration equipment. However, anunderstanding of how operating conditions affect systemdesign, efficiency, energy consumption, and particularly com-pressor performance, should be of great worth to the technician.

Data Points and System Design Calculations:

Refrigeration Effect (RE): This is the total heat transfer, inBtu/lb, from the refrigerated space to the refrigerant. H1 minusH4 (H1 is the enthalpy of the refrigerant at point #1 in Fig. 3,and so forth).

Note: For the purpose of this discussion, H1 will be considered thepoint where the evaporation line intersects with the saturated vaporline. In the real world, the location of H1 would be to the right of thesaturated vapor line, reflecting the superheated vapor at the evapora-tor outlet. With an expansion valve maintaining a typical amount ofsuperheat (in the 4° - 6° range for low temperature applications), theheat transferred to the vapor is minimal (less than 1 Btu/lb). The heattransferred to the vapor between the suction piping outside of therefrigerated space and the compressor cylinder inlet is never consid-ered as part of the refrigeration effect, as this heat was not trans-ferred from the refrigerated space.

Heat of Compression (HOC): This is the amount of heat addedto the refrigerant from the compression process (representedby H2 minus H1 on the chart).

Heat of Rejection (HOR): This is the amount of heat that hasto be rejected at the condenser…the heat transferred to therefrigerant from the refrigerated space (RE), and the heat trans-ferred to the refrigerant during compression (HOC). It is thisvalue, plus some safety factor, from which the condenser selec-tion is made (represented by H2 minus H3 on the chart).

Refrigerant Circulation Rate (RCR): The amount of refrigerantin lbs/min which must circulate in the system to meet thedemands of the load.

(200 Btu/min – ton)RE (Btu/lb)

Compressor Horsepower Required: The horsepower/tonrequirement to meet the load demand. Contrary to popularbelief, 1 horsepower and 1 ton are not necessarily synonymous.

RCR (lb/min-ton) X HOC (Btu/lb) X 778 ft-lb/Btu33,000 ft-lb/min-hp

FORM 5-200 /

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TYPICAL CYCLE - 1Open Drive Compressor40°F Suction S.H. @ Comp. Inlet215°F Discharge Temp.

TYPICAL CYCLE - 1Open Drive Compressor40°F Suction S.H. @ Comp. Inlet215°F Discharge Temp.

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tonton

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Fig. 4 – Typical Cycle #1

Compressor Volume Required: The compressor cylinder vol-ume requirement needed to pump the RCR, in cu ft/ton. Thevapor specific volume is read on the lines of constant volume.

RCR (lb/min-ton) X vapor specific volume (cu ft/lb)

Let’s look at a few typical system scenarios, plot them on theP-H diagram, and then compare the performance measurements.

Typical Cycle #1 (Fig. 4): It is neither realistic nor safe to havea saturated vapor at the compressor inlet. Because liquid can-not be present with superheated vapor, some amount of super-heat at the compressor inlet becomes the margin necessaryto insure the safety of the compressor. Here we see the suc-tion vapor superheated to +20ºF (40º superheat). This is theresult of an expansion valve set to maintain some amount ofsuperheat, plus the temperature increase the refrigerant vaporexperiences in the suction line. The suction line connected tothe compressor will have some accumulation of frost on it.This is the result of the 20°F pipe temperature and mositure inthe air, NOT the result of floodback. Floodback is not possiblewhen the vapor’s temperature is 40° above the saturation tem-perature (40° superheat). Do not confuse frost buildup withfloodback.

The suction superheat will insure that the compressor is pro-tected from liquid flooding. The cost of this protection comesin the form of a larger compressor volume requirement. This isdue to the warmer, thinner suction vapor; at +20ºF the specificvolume of the vapor (measured in cu.ft./lb) is greater than itwas at -20ºF. The compressor’s cylinders are a measured vol-ume...they never change. The density of the refrigerant maychange, and this will affect the pounds per minute of refriger-ant that the compressor will pump, however the volumepumped remains constant. So, because the suction vapor is lessdense, we now require more compressor cylinder volume topump the same mass flow.

While all the various data points and system design calcula-tions are listed, they will appear in a comparative chart later.

Typical Cycle #1A (Fig. 5): Using an open drive compressor,with a 20ºF vapor temperature at the compressor inlet, we seethe benefit of subcooling the liquid to 50ºF. Note the change inrefrigerant quality at the TEV outlet. Instead of 65% liquid, wenow have 80% liquid. Because the difference between liquidtemperature and evaporator temperature has been reduced,there is less refrigerant flashing during the expansion process.As a result, the TEV and distributor nozzle (if used) can possi-bly be downsized.

/ FORM 5-200

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102102 10810840402222 135135

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TYPICAL CYCLE - 1AOpen Drive Compressor40°F Suction S.H. @ Comp. InletLiquid Subcooled to 50°F

TYPICAL CYCLE - 1AOpen Drive Compressor40°F Suction S.H. @ Comp. InletLiquid Subcooled to 50°F

lblb

lblb

lblb

HOR = 95 Btu

HOC = 27 Btu

NRE = 80 Btu

HOR = 95 Btu

HOC = 27 Btu

NRE = 80 Btu

tonton

min - tonmin - ton

tonton







Fig. 5 – Typical Cycle #1A

The use of a liquid to suction heat exchanger will yield sub-cooled liquid, but at the expense of higher suction vapor tem-peratures. While this method of subcooling will insure vaporfree liquid refrigerant at the TEV inlet, it has little effect onincreasing system efficiency. The benefit realized resultingfrom subcooling will be offset by the higher suction vapor tem-peratures, and the volume requirement penalty they impose. Inthis supermarket example the liquid for the low temperaturerack is subcooled using the medium temperature rack. There isno heat gain to the low temperature rack as a byproduct of thesubcooled liquid. The refrigerant is subcooled on a rack that isoperating between 2 to 2-1/2 horsepower per ton, and the ben-efit is being experienced on a rack that is operating near 5horsepower per ton.

The comparative chart will show the benefits: reduced refriger-ant circulation rate, reduced horsepower requirement, andreduced compressor volume requirement. If the subcooling canbe accomplished very inexpensively, such as ambient subcool-ing with a receiver bypass, the energy savings can be enor-mous. For more details on this see Sporlan Form 90-134.

Typical Cycle #2 (Fig. 6): The advantage of using a hermeticcompressor is the elimination of either belts or drive motorcouplings which require precise alignment, and crankshaft

seals. The disadvantage is that there is now an electric motor inthe refrigerant circuit (at least on suction vapor cooled hermet-ic compressors). In addition to the guaranteed system contam-ination problem when a hermetic motor burns, you have theheat from the motor being transferred to the refrigerant vapor.An approximate 80ºF temperature increase can be expectedbetween the vapor entering the compressor service valve, andthe vapor entering the compressor cylinders. This brings thesuction vapor temperature up to 100ºF, with a correspondingincrease in discharge temperature…now approaching 300ºF.This is the upper limit at which most compressor manufacturersagree shouldn’t be exceeded. The refrigerant vapor at 100ºFhas a specific volume approximately 20% higher than at 20ºF.This translates into a compressor volume requirement whichwill be approximately 20% greater. The higher suction vaportemperature also results in a higher HOC, which raises thehorsepower requirement.

Typical Cycle #3 (Fig. 7): Now, let’s take a look at the realworld. It is now the dead of summer – the most extreme condi-tion for the equipment. When the system design and equipmentselection was made, it was this summer weather that was usedas the worst case condition of operation. Now, your companywas so busy in the winter that the yearly preventative mainte-nance was not done. Or perhaps it’s one of those customers

FORM 5-200 /

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SCALE CHANGE

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TYPICAL CYCLE - 2Hermetic Compressor40°F Suction S.H. @ Comp. Inlet295°F Discharge Temp.

TYPICAL CYCLE - 2Hermetic Compressor40°F Suction S.H. @ Comp. Inlet295°F Discharge Temp.

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120°F Suction Superheat 80°F Vapor Temperature increase from hermetic motor windings120°F Suction Superheat 80°F Vapor Temperature increase from hermetic motor windings

lblb

lblb

lblb

HOR = 152 - 40 = 112 Btu

HOC = 152 - 121 = 31 Btu

NRE = 62 Btu

HOR = 152 - 40 = 112 Btu

HOC = 152 - 121 = 31 Btu

NRE = 62 Btu

tonton

min - tonmin - ton

tonton








who never wants to spend money on prevention…the one youtell “you can either pay me a little now, or you can pay me a lotmore, later”. In either case, the condenser is dirty, and the bot-tom line is that the condensing temperature has increased; from100ºF to 120ºF.

With the higher discharge pressure, the compression ratio hasincreased from 8.5:1 to 11:1. This higher pressure that thecompressor is pumping against has a twofold negative impact.First, the motor amperage will be higher. Secondly, the volu-metric efficiency of the compressor will be reduced.

There is a pocket of vapor between the bottom of the valveplate, and the top of the piston, called the clearance volume. Itis there to insure that the piston does not run into the valveplate during the operation of the compressor. The vapor in thispocket requires re-expansion to a lower pressure before anynew suction vapor can enter the compressor’s cylinders.

Let’s take a look at the compression cycle. It begins when suc-tion vapor enters a cylinder as its piston is traveling down. Asthe piston starts to travel back up, reducing the volume of thecylinder, the vapor pressure increases. When the piston reachesthe top of its stroke, the entire volume of compressed vapor willhave exited the cylinder through the discharge valves EXCEPT

for the vapor trapped in the clearance volume. It too is at thedischarge pressure. Before any suction vapor can re-enter thecylinders, the clearance volume vapor must experience a reduc-tion in pressure to a level slightly below that of the suction pres-sure. Otherwise, there would be no flow into the cylinder. It ispiston travel, which increases the cylinder volume, that reducesthis clearance volume pressure. This portion of the piston trav-el, which is entirely dedicated to lowering the clearance volumepressure, performs no useful work at all. In fact, it is advanta-geous to keep this to a minimum. The higher the clearance vol-ume pressure is above the suction pressure, the more of thiswasted piston travel will be required. Simply put, this is the def-inition of a high compression ratio (substituting discharge pres-sure for clearance volume pressure, as they are one and thesame).

Note the higher HOC. Referring to the system design calcula-tions reveals that more horsepower will be required. In addi-tion, the quality of the saturated liquid/vapor mixture at theTEV outlet has further deteriorated, resulting in a lower RE.Referring again to the system design calculations will revealthat the lower the RE is, the higher the RCR has to be. This, inturn, will require more cylinder volume to meet the demand ofan increased RCR. In a typical supermarket there are backupcompressors that only operate under high load conditions, so

/ FORM 5-200

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102102 1211214545 1551554040 152152

320°F Discharge Temp.320°F Discharge Temp.

295°F Discharge Temp.295°F Discharge Temp.

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lblbHOR @ 100°F = 112 Btu

HOC @ 100°F = 31 Btu

NRE @ 100°F Cond. = 62 Btu

HOR @ 100°F = 112 Btu

HOC @ 100°F = 31 Btu


lblb

lblb

HOR @ 120°F = 110 Btu

HOC @ 120°F = 34 Btu


HOR @ 120°F = 110 Btu

HOC @ 120°F = 34 Btu


lblb

lblb

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TYPICAL CYCLE - 3Dirty Condenser - 120°F Cond. TempHermetic Compressor120°F S.H. @ Cylinder Inlet

TYPICAL CYCLE - 3Dirty Condenser - 120°F Cond. TempHermetic Compressor120°F S.H. @ Cylinder Inlet


extra capacity is not a problem. For the store owner, paying forthe additional electricity to operate the extra compressor(s)might be a major problem.

Finally, the discharge temperature has increased to 320ºF. Atthis elevated temperature the mineral oil’s lubrication qualitieswill be diminished. Additionally, at 320ºF mineral oil will mostcertainly start to decompose. Under certain circumstances, therefrigerant (R22) may start to decompose as well. All of thisspells a short destructive life for the compressor.

Typical Cycle #4 (Fig. 8): Now for the double whammy. Notonly is the condenser filthy, but the TEV’s were never adjust-ed. They are operating at abnormally high superheats, which ineffect have reduced the size of the evaporators. As a result, thedischarge air temperature in the glass door frozen food displaycases is too high, causing the frozen juice to melt.

Because it is the dead of summer, and the service call log isoverflowing, a quick solution to this problem would be great.So, in order to compensate for the high discharge air tempera-tures in the fixture lineup mentioned above, the service techni-cian decides to lower the EPR setting. It doesn’t occur to himthat a -20ºF saturated suction temperature should be lowenough to achieve a -10ºF discharge air temperature.

So, this makes sense…if a lower discharge air temperature isrequired, what easier way is there to accomplish it? By reduc-ing the refrigerant pressure in the evaporator, the saturationtemperature will be reduced. This allows for a higher tempera-ture difference between the air entering the evaporator and theheat transfer medium (the refrigerant), resulting in a higher rateof heat transfer…or in simple terms a lower discharge air tem-perature. There is only one problem. The EPR is already wideopen. Hey…how about lowering the set point for the commonsuction pressure on the rack’s energy management control sys-tem. There’s plenty of extra horsepower in the form of idlecompressors, so capacity isn’t an issue…just bring anothercompressor on.

With a simple adjustment, the fixture lineup’s discharge airtemperature can be brought in line. The pressure set point islowered to 20 psia (-29ºF). Simple…yes…but there’s a doozyof a problem with this approach.

Lowering the suction pressure will increase the compressionratio: in this example from 11:1 to 13.6:1…an approximate20% increase over the compression ratio in Typical Cycle #3.This further reduces the compressor volumetric efficiency, andincreases the horsepower requirement.

FORM 5-200 /

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REFRIGERANT 22

SCALE CHANGE

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102102 1241244545 165165

TYPICAL CYCLE - 4Dirty Condenser - 120°F Cond. TempHermetic Compressor130°F S.H. @ Cylinder InletSuction Pressure lowered to 20 psia (-29°F)

TYPICAL CYCLE - 4Dirty Condenser - 120°F Cond. TempHermetic Compressor130°F S.H. @ Cylinder InletSuction Pressure lowered to 20 psia (-29°F)

lblb

lblb

lblb

HOR = 120 Btu

HOC = 41 Btu

NRE = 57 Btu

HOR = 120 Btu

HOC = 41 Btu

NRE = 57 Btu

tonton

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tonton








The lower the suction pressure is, the higher the refrigerant’sspecific volume will be, which translates into a greater com-pressor volume requirement. Simply reducing the suction pres-sure has resulted in a 30% increase in volume requirement overTypical Cycle #3.

The final blow is a higher discharge temperature, which is nowapproaching 370ºF…way beyond any margin of safety. It’sguaranteed that the mineral oil will vaporize off the cylinderwalls. This leaves the compressor’s metal to metal movingparts vulnerable to accelerated wear, and certain failure. A lookin the crankcase will reveal a black sludge where oil used to be.The crankshaft lubrication passages are now in danger of plug-ging up with the oil breakdown. It’s quite possible that some ofthis oil breakdown will end up in TEV ports, which will leavethem unable to control proper superheat. Unless preventativemeasures are taken, this compressor is headed for the scrapheap. In addition, the operation has become extremely ineffi-cient, resulting in even higher energy costs.

Let’s take a look at the chart comparing the system design cal-culations of the various cycles discussed. This can be seen inFig. 9. The important values to look at are:

Compression Ratio: The ratio of absolute dischargepressure/absolute suction pressure. The compressor motoramperage will increase as the compression ratio increases.Also, the higher compression ratio results in reduced compres-sor volumetric efficiency. The net result is that it will take moreenergy to run a less efficient compressor. So, additional com-pressors will have to operate to compensate for this.

Discharge Temperature: A good measure of the relative healthor sickness of the system. Most compressor manufacturershighly suggest the discharge temperature be kept below 300ºF.It is not realistic to put a temperature probe inside the cylinderto monitor temperature. Experience has shown that the dis-charge temperature 6” from the discharge service valve will bebetween 50º - 75º less than the actual discharge temperature. Inaddition to cleaning condensers, discharge temperatures can bereduced by proper expansion valve setting, compressor bodycooling fan motors, good suction line insulation, liquid injec-tion and…keeping the suction pressure as high as possible. Formore details regarding discharge temperatures and liquid injec-tion, refer to Sporlan Form 10-197.

Refrigerant Circulation Rate: The larger this number is, themore compressors will be required in operation to achievecapacity. It would be ideal to search out ways to reduce theamount of refrigerant to be pumped. Anything that affects therefrigerant quality at the TEV outlet will influence the RCR.It’s not how much refrigerant we are feeding to the TEV, buthow much of it remains after the expansion process, for the liq-uid refrigerant remaining is our only medium for transferringheat from the refrigerated space. Higher condensing tempera-tures have a negative affect on this. Notice the great reductionin the RCR where liquid subcooling is employed. Again, if thesubcooling can be accomplished inexpensively, the potentialfor savings is great. It was the increase in discharge pressure(temperature) in Typical Cycle #3 which increased the RCR.

Compressor Horsepower: We can all relate to horsepower.The higher the horsepower requirement, the higher the electricbill will be at the end of the month. The subcooled liquidreduced the horsepower requirement. Because the refrigerantquality was reduced at the outlet of the TEV (more liquid pres-ent), it was used more efficiently. Therefore less of it wasrequired to circulate, which caused a reduction in the horse-power required.

HOC and RCR are the two values which will determine thehorsepower requirement. We see a 20% increase in the horse-power requirement in Typical Cycle #3 (as compared toTypical Cycle #2). This comes from the discharge pressurebeing raised (from 210 psia to 274 psia, an increase in theHOC). Such an easy thing to keep the condenser clean; yet howhuge the impact is, if not done.

In Typical Cycle #4 the suction pressure is lowered from 24psia to 19 psia. This 5 psi reduction in suction pressure resultsin a 20% increase in the system horsepower requirement aswell. The lower suction pressure yields a smaller refrigerationeffect, which results in a higher RCR. It also leads to a higherHOC. Since lowering the suction pressure will increase theHOC and RCR, it is important to keep the suction pressure ashigh as possible.

Compressor Volume: The compressor horsepower and thecylinder volume requirement are independent of one another.We see that from Typical Cycle #1 where there is no suctionsuperheat, to Typical Cycle #2 where there is 40º of superheat,there is an approximate 20% increase in the cylinder volumerequirement. The horsepower remains constant, however. It isthe warmer, thinner suction vapor which drives the volumerequirement up. One can see the tremendous penalty for usinga hermetic compressor. Perhaps the reduced maintenance off-sets the higher power bills. While high suction superheatincreases the vapor specific volume (cu. ft./lb), it is the lowersuction pressure in Typical Cycle #4 which has the greatestimpact on increasing the specific volume.

There is a threefold negative affect from lower suction pressure:

First, it will cause an increase in compression ratio, andthe resulting decrease in volumetric efficiency. Loweringthe suction pressure by a few pounds will result in amuch greater increase to the compression ratio than byraising the discharge pressure the same amount.

Secondly, the refrigerant’s specific volume increaseswith a decrease in pressure. As the specific volumeincreases (a decrease in density), more cylinder volumeis required to pump the same mass flow.

Thirdly, the compression process follows the constantentropy lines. Therefore, the lower the suction pressue is,while following a constant entropy line during compres-sion to a given condensing pressure, the higher the result-ing discharge temperature will be. This has destructiveconsequences on the chemicals in the system, of which inpart the compressor relies on for a long life.

/ FORM 5-200

Comparative Datawith Varying Conditions

Typical Cycle - 2

Typical Cycle - 3Hermetic Compressor (add 80°F S.H.)120°F S.H. vapor entering cylindersDirty Condenser (120°F Cond. Temp)

Hermetic Compressor (add 80°F S.H.)120°F S.H. vapor entering cylindersDirty Condenser - 120°F Cond. TempSuction pressure lowered tocompensate for starving TEV's

Typical Cycle - 4

100 120 365 13.9 245 41 120 57 1.39 3.39 3.54 12.42

100 120 320 11.1 200 34 110 57 1.68 2.81 2.72 9.54

Suct

ion

Tem

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ense

rTe

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(°F) (°F) (°F) (°F) Btu/lb Btu/lb Btu/lb hp/ton cu ft/lb cfm/ton

Ideal Cycle -20 100 180 8.5 80 27 89 62 2.30 2.05 2.07 6.68

Typ ical Cycle - 120 100 215 8.5 115 27 95 62 2.30 2.05 2.29 7.39Open Drive Compressor

(40°F Suction S.H.)

Typ ical Cycle - 1A

20 100 215 8.5 115 27 95 80 2.96 1.59 2.29 5.73Open Drive Compressor(40°F Suction S.H.)Liquid subcooled to 50°F

-30

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3.23 2.36 2.72 8.77Hermetic Compressor (add 80°F S.H.)120°F S.H. vapor entering cylinders

REFRIGERANT 22

Fig. 9 – Comparison of System Design Calculations

FORM 5-200 /

The simplest way to insure against abnormally low suctionpressure is to set the expansion valve for the proper superheat.When the TEV is set properly, the evaporator is efficiently usedas a heat transfer device due to the boiling liquid presentthrough most of its tubing length. High superheat settingsreduce the amount of liquid available for heat transfer, in effectreducing the evaporator capacity. This can be negated by low-ering the suction pressure, which lowers the saturation temper-ature, and yields a higher temperature difference between theentering air and the refrigerant. Evaporator capacities increaseas the TD increases, so this allows for proper product temper-ature. It is done at the expense of lower suction pressure, andthe negative impact it has on system health.

While several “realistic” system scenarios have been plottedand dissected, as stated earlier there are still some aspects ofeach cycle that are represented in an ideal fashion. There arepressure losses in the tubing, valves, accessories, etc. which arenot shown. The compression process occurs at a constantentropy ONLY in the ideal cycle. In the real world, entropywill increase during the compression process, resulting in evenhigher discharge temperatures and HOC values. These factorsshould not detract from the basic focus of this article: to giveone a solid foundation of the refrigeration cycle, based on whathappens to the refrigerant as it travels throughout the system,and how the operating conditions can influence the relativehealth and efficiency of the system.

5-200

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